Bearing housing shroud

ABSTRACT

A turbocharger including a turbine wheel having a hub-to-tip ratio of no more than 60% and blades with a high turning angle, a turbine housing forming an inwardly spiraling primary-scroll passageway that significantly converges to produce highly accelerated airflow into the turbine at high circumferential angles, and a two-sided parallel compressor. The compressor and turbine each produce substantially no axial force, allowing the use of minimal axial thrust bearings. The bearing housing forms a shroud for the internal side of the compressor wheel.

This application is a continuation-in-part of application Ser. No.12/799,182, filed Apr. 19, 2010, which is incorporated herein byreference for all purposes.

The present invention relates generally to turbochargers and, moreparticularly, to a bearing housing forming a shroud for a two-sidedradial compressor.

BACKGROUND OF THE INVENTION

With reference to FIG. 1, a typical turbocharger 101 having a radialturbine includes a turbocharger housing and a rotor configured to rotatewithin the turbocharger housing along an axis of rotor rotation 103 onthrust bearings and two sets of journal bearings (one for eachrespective rotor wheel), or alternatively, other similarly supportivebearings. The turbocharger housing includes a turbine housing 105, acompressor housing 107, and a bearing housing 109 (i.e., a centerhousing that contains the bearings) that connects the turbine housing tothe compressor housing. The rotor includes a turbine wheel 111 locatedsubstantially within the turbine housing, a compressor wheel 113 locatedsubstantially within the compressor housing, and a shaft 115 extendingalong the axis of rotor rotation, through the bearing housing, toconnect the turbine wheel to the compressor wheel.

The turbine housing 105 and turbine wheel 111 form a turbine configuredto circumferentially receive a high-pressure and high-temperatureexhaust gas stream 121 from an engine, e.g., from an exhaust manifold123 of an internal combustion engine 125. The turbine wheel (and thusthe rotor) is driven in rotation around the axis of rotor rotation 103by the high-pressure and high-temperature exhaust gas stream, whichbecomes a lower-pressure and lower-temperature exhaust gas stream 127and is axially released into an exhaust system (not shown).

The compressor housing 107 and compressor wheel 113 form a compressorstage. The compressor wheel, being driven in rotation by the exhaust-gasdriven turbine wheel 111, is configured to compress axially receivedinput air (e.g., ambient air 131, or already-pressurized air from aprevious-stage in a multi-stage compressor) into a pressurized airstream 133 that is ejected circumferentially from the compressor. Due tothe compression process, the pressurized air stream is characterized byan increased temperature over that of the input air.

Optionally, the pressurized air stream may be channeled through aconvectively cooled charge air cooler 135 configured to dissipate heatfrom the pressurized air stream, increasing its density. The resultingcooled and pressurized output air stream 137 is channeled into an intakemanifold 139 on the internal combustion engine, or alternatively, into asubsequent-stage, in-series compressor. The operation of the system iscontrolled by an ECU 151 (engine control unit) that connects to theremainder of the system via communication connections 153.

U.S. Pat. No. 4,850,820, dated Jul. 25, 1989, which is incorporatedherein by reference for all purposes, discloses a turbocharger similarto that of FIG. 1, but which has an axial turbine. The axial turbineinherently has a lower moment of inertia, reducing the amount of energyrequired to accelerate the turbine. As can be seen in FIG. 2, theturbine has a scroll that circumferentially receives exhaust gas at theradius of the turbine blades and (with reference to FIG. 1) axiallyrestricts the flow to transition it to axial flow. It thus impacts theleading edge of the turbine blades in a generally axial direction (withreference to col. 2).

For many turbine sizes of interest, axial turbines typically operate athigher mass flows and lower expansion ratios than comparable radialturbines. While conventional axial turbines generally offer a lowerinertia, albeit with some loss of efficiency and performance, theysuffer from an inability to be efficiently manufactured in the smallsizes usable with many modern internal combustion engines. This is,e.g., due to the exceptionally tight tolerances that would be required,due to aerodynamic limitations, and/or due to dimensional limitations oncreating small cast parts. Axial turbines also lack the ability toperform well at higher expansion ratios, such as are typically neededdue to the pulsing nature of the exhaust of an internal combustionengine. Furthermore, conventional axial turbines have a significantchange in static pressure across the blades, causing significant thrustloads on the thrust bearings of the rotor, and potentially causingblowby.

In some conventional turbochargers the turbines and compressors areconfigured to exert axial loads in opposite directions so as to lessenthe average axial loads that must be carried by the bearings.Nevertheless, the axial loads from the turbines and compressors do notvary evenly with one another and may be at significantly differentlevels, so the thrust bearings must be designed for the largest loadcondition that may occur during turbocharger use. Bearings configured tosupport high axial loads waste more energy than comparable low-loadbearings, and thus turbochargers that must support higher axial loadslose more energy to their bearings.

Accordingly, there has existed a need for a turbocharger turbine havinga low moment of inertia, and characterized by a small size that does notrequire exceptionally tight tolerances, while having reasonableefficiency both at both lower and higher expansion ratios, and smalleraxial loads. Preferred embodiments of the present invention satisfythese and other needs, and provide further related advantages.

SUMMARY OF THE INVENTION

In various embodiments, the present invention solves some or all of theneeds mentioned above, typically providing a cost effective turbochargerturbine characterized by a low moment of inertia, and having a smallsize that does not require exceptionally tight tolerances, whileoperating at reasonable efficiency levels at both at both lower andhigher expansion ratios, and having only small changes in static loads.

A typical turbocharger under the invention is provided with aturbocharger housing including a turbine housing, a bearing housing thatis integral, and a compressor housing. The turbine housing is axiallyconnected to the bearing housing, and the bearing housing is axiallyconnected to the compressor housing. The turbocharger further includes arotor configured to rotate within the turbocharger housing along anaxial axis of rotor rotation, the rotor being supported by bearinghousing bearings. The rotor includes a turbine wheel, a two-sidedcompressor wheel, and a shaft extending axially along the axis of rotorrotation and connecting the turbine wheel to the compressor wheel. Thebearings constrain the rotor from movement other than axial rotation(i.e., rotation around the axis of rotor rotation). As a two-sidedcompressor wheel, the compressor wheel includes a first set ofcompressor blades facing axially away from (and being farther from) theturbine wheel and a second set of compressor blades facing axiallytoward (and being closer to) the turbine wheel. The compressor housingforms a first shroud wall that substantially conforms to a space throughwhich the first set of compressor blades is configured to rotate,leaving only a small gap there-between (i.e., between the blades and theshroud wall).

The bearing housing forms a second shroud wall that substantiallyconforms to a space through which the second set of compressor blades isconfigured to rotate, leaving only a small gap there-between.Advantageously, having the bearing housing form the second shroud wallavoids complications suffered by previously known structures that have aseparate insert, intermediate the compressor housing and the bearinghousing, to form a second shroud wall. With the present structure, fewerparts are required, and as a result there is less cost and lesstolerance buildup issues (i.e., there are fewer serially connected partsthat each contribute to the tolerances of the shrouds.

In another feature of the invention, the bearing housing sequentiallyincludes a turbine-end portion, a central portion, one or more supports,and a compressor-end portion. The turbine-end portion and centralportion forms structures there-between that support the bearings. Thecompressor end portion forms the second shroud wall. The one or moresupports extend between the central portion and the compressor-endportion. The one or more supports are typically a plurality of supportsthat are radially separated from, and circumferentially spaced around,the axis of rotor rotation.

Advantageously, the plurality of supports forms a plurality of radialopenings into the bearing housing between the compressor-end portion andthe central portion. These openings form an inlet passage that providesfor airflow leading to an inducer of the second set of blades.

Other features and advantages of the invention will become apparent fromthe following detailed description of the preferred embodiments, takenwith the accompanying drawings, which illustrate, by way of example, theprinciples of the invention. The detailed description of particularpreferred embodiments, as set out below to enable one to build and usean embodiment of the invention, are not intended to limit the enumeratedclaims, but rather, they are intended to serve as particular examples ofthe claimed invention.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a system view of a prior art turbocharged internal combustionengine.

FIG. 2 is a cross-sectional plan view of a turbocharger embodying thepresent invention.

FIG. 3 is a cross-sectional side view of the turbocharger depicted inFIG. 2, taken along line A-A of FIG. 2.

FIG. 4 is a plan view of certain critical flow locations relative to aturbine wheel depicted in FIG. 2.

FIG. 5 is a depiction of the camber of a turbine blade depicted in FIG.2.

FIG. 6 is a perspective view of the turbine wheel depicted in FIG. 2.

FIG. 7 is a perspective view of the bearing housing, the turbinehousing, and the compressor wheel depicted in FIG. 2, with thecompressor housing removed.

FIG. 8 is a perspective view of the bearing housing depicted in FIG. 7.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

The invention summarized above and defined by the enumerated claims maybe better understood by referring to the following detailed description,which should be read with the accompanying drawings. This detaileddescription of particular preferred embodiments of the invention, setout below to enable one to build and use particular implementations ofthe invention, is not intended to limit the enumerated claims, butrather, it is intended to provide particular examples of them.

Typical embodiments of the present invention reside in a motor vehicleequipped with a gasoline powered internal combustion engine (“ICE”) anda turbocharger. The turbocharger is equipped with a unique combinationof features that may, in various embodiments, provide the aerodynamicbenefits of a zero reaction turbine with the geometric benefits of afifty percent reaction turbine, and/or provide significantly improvedsystem efficiencies by combining less efficient components in a mannerthat reduces the bearing requirements, and thereby forms a system with ahigher efficiency than the comparable unimproved system.

The turbine is configured to operate at reasonable efficiency levels atboth lower and higher expansion ratios, having only small changes instatic pressure across the turbine wheel (and thereby low rotor thrustloads), while it has a low moment of inertia, and is characterized by asmall size, but does not require exceptionally tight tolerances. Incombination with this, the compressor is also characterized by low axialthrust loads, providing for the turbocharger to require a thrust bearingthat is significantly more efficient than is used in comparableconventional turbochargers.

With reference to FIGS. 2 & 3, in a first embodiment of the invention atypical internal combustion engine and ECU (and optionally anintercooler), such as are depicted in FIG. 1, are provided with aturbocharger 201 that includes a turbocharger housing and a rotorconfigured to rotate within the turbocharger housing along an axis ofrotor rotation 203 on a set of bearings. The turbocharger housingincludes a turbine housing 205, a compressor housing 207, and a bearinghousing 209 containing bearings 210 (i.e., a center housing thatcontains radial and thrust bearings) that connects the turbine housingto the compressor housing. The rotor includes an axial turbine wheel 211located substantially within the turbine housing, a radial compressorwheel 213 located substantially within the compressor housing, and ashaft 215 extending along the axis of rotor rotation, through thebearing housing, to connect the turbine wheel to the compressor wheeland provide for the turbine wheel to drive the compressor wheel inrotation around the axis of rotation.

The turbine housing 205 and turbine wheel 211 form a turbine configuredto circumferentially receive a high-pressure and high-temperatureexhaust gas stream from an exhaust manifold of the engine (such as theexhaust gas stream 121 from the exhaust gas manifold 123, as depicted inFIG. 1). The turbine wheel (and thus the rotor) is driven in rotationaround the axis of rotor rotation 203 by the high-pressure andhigh-temperature exhaust gas stream acting on a plurality of blades 231of the turbine wheel. The exhaust gas stream becomes a lower totalpressure exhaust gas stream while passing through the blades, and issubsequently axially released via a turbine outlet 227 into an exhaustsystem (not shown).

The compressor housing 207 and compressor wheel 213 form a radialcompressor. The compressor wheel, being driven in rotation by theexhaust-gas driven turbine wheel 211 (via the shaft 215), is configuredto compress axially received input air (e.g., ambient air, oralready-pressurized air from a previous-stage in a multi-stagecompressor) into a pressurized air stream that may be ejectedcircumferentially from the compressor and sent on to an engine inlet(such as pressurized air stream 133 that is sent on to the engine inlet139, as depicted in FIG. 1).

Turbine Volute

The turbine housing 205 forms an exhaust gas entrance passageway 217leading into a primary-scroll passageway 219 configured to receive theexhaust gas stream from the engine in a direction normal to and radiallyoffset from the rotor axis of rotation 203. The primary-scrollpassageway forms a spiral adapted to significantly accelerate the speedof the gas stream to a high speed, which may be a supersonic speed forat least some operating conditions of the turbine (and its relatedengine). More particularly, the primary-scroll passageway turns theexhaust gas both inwardly around the axis of rotation 203 and axiallytoward the axial turbine wheel 211, thereby achieving (for some standardoperating conditions of the engine) a supersonic flow having both adownstream axial component 221 and a downstream circumferentialcomponent 223.

Effectively, this configuration takes advantage of the conservation ofangular momentum (rather than a convergent divergent nozzle) to achievea high-speed airflow that may include a shockless transition tosupersonic speeds for at least some operating conditions. Typically, aspiral characterized by a large radius change is required to achievethis change in velocity, and even though the resulting airstream isturned axially into an axial turbine wheel, it has a very high-speedcircumferential component.

This circumferential component is achieved without the use of turningvanes, which would cause additional losses. Thus, the turbine inlet ofthis embodiment is of a vaneless design. As compared to a design withvanes, such a design advantageously is cost efficient, reliable (in thatit eliminates parts from an environment in which they are likely toerode), avoids friction pressure losses, and avoids establishing acritical throat area that could choke the flow in some operatingconditions.

With reference to FIGS. 2-4, this potentially supersonic flow of theaccelerated exhaust gas stream in the inner radius of the primary-scrollpassageway is directed into the turbine wheel 211. More particularly,the primary-scroll passageway is an inwardly spiraling passagewaycharacterized by a primary-scroll inlet port 225 that connects theprimary-scroll passageway to the exhaust gas entrance passageway 217.The primary-scroll passageway substantially forms a convergentpassageway that spirals inward enough and converges enough to acceleratethe exhaust gas, and to achieve supersonic speeds for at least somestandard operating conditions of the engine (and thus of theturbocharger) as the exhaust gas turns axially downstream and impingeson the axially upstream end 233 of the blades 231.

The primary-scroll inlet port 225 is a planar location located along thepassageways within the turbine that the exhaust gas travels throughprior to reaching the turbine wheel. The location of the primary-scrollinlet port is defined relative to an opening in the passageway, which ischaracterized by a tongue-like shape when viewed in a cross-sectiontaken normal to the rotor axis of rotation 203.

More particularly, the structure of a tongue 235 appears as a protrusionhaving a tip when viewed in the cross-section of FIG. 3. It should benoted that in some embodiments this structure will not vary in shapewhen the cross section is taken at different axial locations. In otherembodiments the structure forming the tongue 235 may be shaped such thatthe location of the tip of the tongue varies when viewed incross-sections taken at different axial locations.

The primary-scroll inlet port 225 is located at the tip of the tongue235. To any extent that the circumferential location of the tip of thetongue appears to vary with the axial location of the cross-sectionconsidered, the primary-scroll inlet port 225 is defined to be at themost upstream location of the tip of the tongue, i.e., the upstream-mostlocation at which the housing opens such that it is no longer radiallyinterposed between the exhaust gas stream and the blades (even thoughthe blades are axially offset from the exhaust gas stream). For thepurposes of this application, the primary-scroll inlet port 225 isdefined as the smallest planar opening from the exhaust gas entrancepassageway 217 into the primary-scroll passageway 219, at the tip of thetongue. In other words, it is at the downstream end of the exhaust gasentrance passageway at the location at which the stream opens up to theblades.

The primary-scroll passageway 219 starts at the primary-scroll inletport 225, and spirals inward 360 degrees around the axis of rotation toform a converging loop that rejoins flow coming in the primary-scrollinlet port 225. This convergent loop accelerates the exhaust gascircumferentially and turns it axially. Throughout the 360 degrees ofthe primary-scroll passageway 219, the accelerated and turned exhaustgas stream impinges on the blades 231, passing between the blades anddriving the turbine wheel 211 in rotation.

In summary, the housing for the axial turbine wheel forms an inwardlyspiraling primary-scroll passageway that surrounds the axis of rotorrotation. It begins at a primary-scroll inlet port 225 that issubstantially radially external to the axially upstream ends of theblades, providing for the passageway to spiral inwardly and turn axiallyto accelerate the exhaust gas flow into the upstream ends of the axialturbine wheel blades.

Corrected Mass Flow

To provide for an adequate level of acceleration of the exhaust gasunder the invention, the primary-scroll passageway 219 is configuredwith sizing parameters such that the corrected mass flow rate surfacedensity of the turbine, when operated at a critical expansion ratio(E_(cr)), exceeds a critical configuration parameter, i.e., a criticalcorrected mass flow rate surface density (D_(cr)). More particularly,the sizing parameters for the scroll include a primary-scroll radiusratio (r_(r)) and a primary-scroll inlet port area (a_(i)), and areselected such that the corrected mass flow rate surface density of theturbine exceeds the critical configuration parameter D_(cr) when theturbine is operated at the critical expansion ratio E_(cr). These sizingparameters are defined relative to the primary-scroll inlet port 225,which is characterized by a centroid 237. For the gas to be axiallyadequately accelerated, this centroid will be substantially radiallyexternal to, and typically axially upstream of, an axially upstream end233 of each blade 231.

The values of some of the above-recited terms are dependent upon thetype of exhaust stream gas that will be driving the turbine. Thisexhaust-stream gas will be characterized by a Boltzmann Constant (k),and by a Gas Constant R-specific (R_(sp)). These constants vary by gastype, but for most gasoline powered engine exhaust gasses, thedifference is anticipated to be small, with the constants beingtypically be on the order of k=1.3 and R_(sp)=290.8 J/kg/K.

The turbine housing has an ability to accelerate the exhaust gas that ischaracterized by the two sizing parameters recited above. The firstsizing parameter, being the primary-scroll radius ratio r_(r) is definedto be a radius of a point 239 at the hub at leading edge of the turbineblades 231 (i.e., at the inner edge of the rotor inlet), divided by aradius of the centroid 237 of the planar area of the primary-scrollinlet port 225. The second, being the primary-scroll inlet port areaa_(i) is defined to be the area of the primary-scroll inlet port 225.

As mentioned above, the geometry of this embodiment of a turbine isdefined relative to operational parameters at the critical expansionratio E_(cr). This critical expansion ratio is obtained from the formula

$E_{cr} = ( \frac{k + 1}{2} )^{(\frac{k}{k - 1})}$

and is a function of the gas-specific Boltzmann's Constant k. A typicalvalue for E_(cr) is 1.832.

As recited above, the dimensions of the primary-scroll passageway 219 ofthis embodiment are limited by a primary-scroll radius ratio r_(r) and aprimary-scroll inlet port area a_(i) that cause the corrected mass flowrate surface density of the turbine to exceed the critical correctedmass flow rate surface density D_(cr). This critical corrected mass flowrate surface density is obtained from the formula

$D_{cr} = {r_{r}\frac{101325}{\sqrt{288R_{sp}}}( {1 - \frac{( {k - 1} )( r_{r} )^{2}}{( {k + 1} )}} )^{(\frac{1}{k - 1})}\sqrt{\frac{2k}{k + 1}}}$

which varies with the primary-scroll radius ratio r_(r).

For any given turbine, exactly one steady-state inlet condition for agiven outlet static pressure (i.e., one inlet total pressure) will drivethe turbine at a given expansion ratio such as the critical expansionratio E_(cr). A variation in the geometry of the volute, e.g., avariation of the radius ratio r_(r) and/or the primary-scroll inlet portarea a, can vary the steady-state mass flow rate that will drive theturbine at the given critical expansion ratio, and thus will affect therelated corrected mass flow rate surface density.

If the primary-scroll radius ratio and the primary-scroll inlet portarea are adequately selected, it will cause the corrected mass flow ratesurface density at the primary-scroll inlet port 225 when driven at thecritical expansion ratio E_(cr) to be greater than the criticalcorrected mass flow rate surface density D_(cr). While the relationshipsbetween the primary-scroll radius ratio, the primary-scroll inlet portarea and the corrected mass flow rate surface density at theprimary-scroll inlet port are complicated, and while they will typicallybe explored experimentally, it may be noted that in general, a higherradius ratio for the same port area will lead to a higher corrected massflow rate surface density.

In an iterative method of designing a turbine under the invention, aperson skilled in the art can first select a composition of an exhaustgas to be received from an engine, look up (from existing sources of gasproperties) the related Boltzmann's Constant k and Gas Constant R_(sp),and calculate the critical expansion ratio E_(cr).

A first configuration of a turbine is then designed. The turbineincludes a volute as described above, with an inwardly spiralingpassageway that turns from a tangential direction to an axial direction,and an axial turbine wheel. The design is characterized by a firstprimary-scroll radius ratio r_(r1) and a first primary-scroll inlet portarea a_(i1).

A prototype is built, put on a gas stand, and run using the selectedexhaust gas. The input total pressure is increased until a calculatedexpansion ratio reaches the critical expansion ratio E_(cr). Thisexpansion ratio is calculated from the total pressure at the inlet andthe static pressure at the outlet. A steady state mass flow rate m, atotal turbine inlet temperature T, and a total inlet pressure p_(i) aremeasured.

The corrected mass flow rate surface density is calculated from themeasured data using the following formula:

$D_{ca} = \frac{m \times \sqrt{\frac{T}{288}}}{\frac{p_{i} \times a_{i}}{101325}}$

where a_(i) is the inlet port area. This calculated corrected mass flowrate surface density D_(ca) is compared to the critical corrected massflow rate surface density D_(cr), which is calculated using thepreviously-identified formula. If the corrected mass flow rate surfacedensity exceeds or equals the critical corrected mass flow rate surfacedensity, then the design of an embodiment of the invention is complete.If the corrected mass flow rate surface density is less than thecritical corrected mass flow rate surface density, then the design isconsidered insufficient to create the high-speed circumferential airflowneeded under the invention, and another iteration of the design andtesting steps are completed.

In this next iteration, the primary-scroll radius ratio r_(r) and/or theprimary-scroll inlet port area a_(i) are appropriately adjusted (e.g.,reduced) to increase the corrected mass flow rate surface density whentaken at the critical expansion ratio E_(cr). This process is repeateduntil a design is found in which the corrected mass flow rate surfacedensity exceeds or equals the critical corrected mass flow rate surfacedensity when taken at the critical expansion ratio E_(cr).

In a potential alternative decision-making process for the above recitediterative design method, the decision to change one or both of thesizing parameters r_(r) and a_(i) is based on testing the axial loadingof the turbine wheel (or the static pressure ratios that cause axialloading) by the exhaust gasses, over critical operating conditions(i.e., conditions that cause operation to occur at the criticalexpansion ratio E_(cr)). Another iteration is conducted if the axialforces are not below a threshold, such as the loading condition when thestatic pressure upstream of the wheel near the wheel hub is greater than120% of the turbine outlet static pressure, i.e. the pressures differ atthe most by 20% of the outlet pressure.

Wheel Blades

With reference to FIGS. 3-5, relative to the downstream axial flowcomponent 221 and downstream circumferential flow component 223, eachblade 231 is characterized by a lower surface 241 (i.e., the surfacegenerally facing circumferentially into the downstream circumferentialflow component) and an upper surface 243 (i.e., the surface generallyfacing circumferentially away from the downstream circumferentialcomponent).

The lower and upper surfaces of the blade 231 meet at a leading edge 245(i.e., the upstream edge of the blade) and a trailing edge 247 (i.e.,the downstream edge of the blade). The blades extend radially outwardfrom a central hub 271 in a cantilevered configuration. They attach tothe hub along a radially inner hub end 273 of the blade, and extend to aradially outer tip end 275 of the blade. The hub end of the bladeextends from an inner, hub end of the leading edge to an inner, hub endof the trailing edge. The tip end of the blade extends from an outer,tip end of the leading edge to an outer, tip end of the trailing edge.

Typical axial turbines are typically provided with blades having bladelengths that are very small compared to the radius of the respectivehub. Contrary to this typical convention, the present embodiment isprovided with blades having a hub-to-tip ratio of less than or equal to0.6 (i.e., the radius of the inner, hub end of the trailing edge is nomore than 60% of the radius of the outer, tip end of the trailing edge).

While convention axial blades having high hub-to-tip ratios also requirelarge numbers of blades to extract any significant amount of energy fromthe exhaust, the present blades are capable of extracting a very highpercentage of the dynamic pressure of the high-speed highly tangentialflow entering the turbine wheel. They can do so with a relativelylimited number of blades, thereby limiting the rotational moment ofinertia of the turbine wheel, and therefore providing for fast transientresponse time. Under numerous embodiments of the invention there are 20or fewer blades, and for many of those embodiments there are 16 or fewerblades.

At any given radial location along the blade, the lower and uppersurfaces are each characterized by a camber, and the blade ischaracterized by a median camber, which for the purposes of thisapplication will be defined as a median camber curve 249 extending fromthe leading edge to the trailing edge at a median location equallybetween the upper and lower surfaces, wherein the median location istaken along a lines 251 extending from the upper camber to the lowercamber, normal to the curve 249 along the median camber curve.

The median camber curve 249 comes to a first end at the leading edge245. The direction of the median camber curve at the leading edgedefines a leading-edge direction 253, and is characterized by aleading-edge direction angle β₁ (i.e., a β₁ blade angle) that is theangular offset between the leading-edge direction and a line that isparallel to the axis of rotation and passing through the leading edge(at the same radial location as the median camber), and therefore alsoto the downstream axial component 221 of the supersonic flow. The β₁blade angle is positive when the leading edge turns in to thecircumferential flow component 223 (as depicted in FIG. 5), and zerowhen the leading edge faces directly along the axial flow component 221.The β₁ blade angle can vary over the radial extent of the leading edge.

The median camber curve 249 comes to a second end at the trailing edge247. The direction of the median camber curve at the trailing edgedefines a trailing-edge direction 255, and is characterized by atrailing-edge direction angle β₂ (i.e., a β₂ blade angle) that is theangular offset between the trailing-edge direction and a line that isparallel to the axis of rotation and passing through the trailing edge(at the same radial location as the median camber). The β₂ blade angleis positive when the trailing edge turns in to the circumferential flowcomponent 223 (as depicted in FIG. 5), and zero when the trailing edgefaces directly along the axial flow component 221. The blade angle β₂can vary over the radial extent of the trailing edge.

The sum of the β₁ and β₂ blade angles at a given radial location on ablade defines a turning angle for the blade at that radial location. Theβ₁+β₂ turning angle can vary over the radial extent of the blade.

While the primary scroll efficiently accelerates the exhaust gas streamand thereby provides for a substantial increase in the dynamic pressureof the exhaust gas stream, it does not typically produce a flow with ahigh degree of axial uniformity, as might be seen from a vaned nozzle.The blades of the present embodiment, and particularly the shapes oftheir leading edges, are tailored so that each radial portion of theblade is best adapted to the flow that occurs at its radial location.This type of tailoring is not typical for conventional axial turbines,as they typically have vaned nozzles providing a high level of flowuniformity, and as they have a much higher hub-to-tip ratio that limitspossible variations between the hub and tip flows.

Under the present embodiment, over the majority of the leading edge ofeach blade, the blade angle faces circumferentially upstream withrespect to the axis of rotation (i.e., the β₁ blade angle is positive).Moreover, the β₁ blade angle is greater than or equal to 20 degrees (andpossibly greater than or equal to 30 degrees) at both the hub end of theleading edge and the mid-span of the leading edge (i.e., the leadingedge half way between its hub end and its shroud end). At the shroud endof the leading edge, the β₁ blade angle is greater than or equal to −20degrees (and possibly greater than or equal to −5 degrees).

Additionally, under the present embodiment, over the majority of theradial extent of each blade, the β₁+β₂ turning angle is positive.Moreover, the turning angle is greater than or equal to 45 degrees atthe hub end of each blade. The turning angle is greater than or equal to80 degrees at the mid-span of each blade. At the shroud end of eachblade, the β₁+β₂ turning angle is greater than or equal to 45 degrees.

The chord line 261 (i.e., the line connecting the leading and trailingedge) has a positive angle of attack with respect to the downstreamaxial component 221, i.e., even though the leading-edge direction facescircumferentially upstream with respect to the axis of rotation, thechord line itself is angled circumferentially downstream with respect tothe axis of rotation. In other words, the leading edge iscircumferentially downstream of the trailing edge. This may vary inother embodiments.

The lower surface 241 of the blade of this embodiment is configured tobe concave over substantially the full chord of the blade. Moreover, atthe majority of radial locations, the lower surface is curved such thatit has a range of locations 263 that are circumferentially downstream ofboth the leading edge and the trailing edge.

Static Pressure Drop

A key feature of the present embodiment of the invention is that itprovides the inertial advantages of a typical axial turbine wheel(having a lower rotational moment of inertia than an equivalent radialturbine wheel), while it greatly enhances the ability of the axialturbine to extract the energy of the exhaust gas stream. To accomplishthis, as previously suggested, the present embodiment is provided with avolute that uses conservation of angular momentum to efficientlyaccelerate the exhaust gas stream and convert a significant portion ofthe total pressure in the exhaust gas stream from static pressure todynamic pressure, and further to provide the accelerated exhaust gasstream to an axial turbine wheel at a significant angle.

The turbine blade is configured to extract a significant portion of theenergy of the dynamic pressure from the flow, but not to significantlychange the static pressure of the flow. As a result of the voluteconverting a significant portion of the static pressure to dynamicpressure, and of the wheel extracting most of the dynamic pressurewithout changing the static pressure of the airstream, the turbineextracts a large percentage of the energy in the exhaust gas streamwithout receiving a significant axial load. A typical embodiment of theinvention will be characterized by a static pressure change across theturbine wheel blades of less than ±20% of the static outlet turbinepressure across the turbine for at least some operating conditions ofthe range of standard operating conditions, thereby causing very littleaxial force to be applied to the turbine wheel. More particularly, theturbine is configured to limit the static pressure upstream of the wheelnear the wheel hub to a value that is not greater than 120% of theturbine outlet static pressure, i.e. the pressures differ at the most by20% of the outlet pressure. Some embodiments of the invention arecharacterized by substantially no static pressure drops across therotor, thereby causing only a negligible axial force on the turbinewheel.

Wheel Hub

With reference to FIGS. 5 & 6, the radial size of the turbine wheel hub271 varies along the blade inner hub end 273 from the leading edge 245of each blade 231 to the trailing edge 247 of each blade, and it isuniform around the circumference. More particularly, the hub is radiallylarger at the leading edge than it is at the trailing edge, and the hubis radially larger at an intermediate axial location between the leadingedge and the trailing edge than it is at either the leading edge or thetrailing edge. This increase in thickness forms a smoothly continuoushump 277 that is axially close to the range of locations 263 on theblade lower surface 241 that are circumferentially downstream of boththe leading edge and the trailing edge (i.e., where the median camber isparallel to the axial component of the flow).

The hump 277 is provided in a location in which significant diffusionoccurs, and it prevents the diffusion from exceeding a critical level atwhich flow separation might occur. The potential for this problem isuniquely substantial because of the unique size and shape of the bladesand the high level of kinetic energy of the flow. Because use of thehump helps avoid flow separation, the hump provides for improvedefficiency over a similar wheel that lacks the hump.

Axially Balanced Compressor

With reference to FIG. 2, the compressor housing 207 and compressorwheel 213 form a dual, parallel, radial compressor. More particularly,the compressor wheel has back-to-back oriented impeller blades. A first,external set of compressor impeller blades 301 are oriented in aconventional configuration with an inlet facing axially outward (awayfrom the turbine wheel) to receive air from that direction. A second,internal set of compressor impeller blades 303 are oriented in a reverseconfiguration with an inlet facing axially inward (toward the turbinewheel) to receive air brought in tangentially and turned to travelaxially into the second set of impeller blades. The second set ofcompressor blades are closer to the turbine wheel than the first set ofcompressor blades. The first and second set of compressor impellerblades can be manufactured in the form of a single, integral wheel,e.g., as illustrated, or may comprise an assembly of a plurality ofparts.

The compressor housing 207 is configured to direct inlet air to each setof compressor blades in parallel, and to direct the passage ofpressurized gas from each compressor. In this embodiment, the compressorhousing is configured to partially form two separate axially positionedair inlets; namely, a first air inlet passage 305, that is positionedadjacent an end of the compressor housing to pass inlet air in an axialdirection to the inlet of the first set of compressor blades 301, and asecond air inlet passage 307 that is separate from the first air inletpassage 305. The second air inlet passage is configured to pass inletair in first a radial and then an axial direction to the inlet of thesecond set of compressor blades 303. Pressurized air that is provided bythe compressor wheel 213 is directed radially from each set of impellerblades 301 and 303 through a single diffuser passage 311 to a compressorvolute 313.

This dual-path, parallel, radial compressor configuration, whiletypically being less efficient than a comparable single-path radialcompressor, will operate at higher speeds and might producesubstantially no axial loading in steady state operation. The higheroperating speeds will typically better match the operational speeds ofthe axial turbine.

Bearing Housing

With reference to FIGS. 2, 7 and 8, the bearing housing 209 is anintegral body that axially includes a turbine-end portion 401, a centralportion 403 and a compressor-end portion 405. Between the turbine-endportion and central portion the bearing housing houses bearings, i.e.,they form structures that contain and support the journal and thrustbearings, which carry the shaft 215 of the rotor.

The compressor-end portion 405 of the bearing housing 209 is held inplace with respect to the remainder of the bearing housing by one ormore, and preferably a plurality (e.g., three) of axially extendingbearing-housing supports 406 that extend between, and are integral with,the compressor-end portion and the central portion 403. These supportsare radially separated from, and circumferentially spaced around, theaxis of rotor rotation 203, thereby forming a plurality of radialopenings into the bearing housing between the compressor-end portion andthe central portion.

The bearing housing central portion 403 forms a first wall 407 of thesecond inlet passage 307 of the compressor. A second wall 409 of thesecond inlet passage is formed by an internal face of the compressor-endportion 405 of the bearing housing 209. The second inlet passage extendsradially in through the plurality of radial openings between thesupports 406 to a second inlet portion 408 of the compressor housing.

A first diffuser wall 411 is formed by an internal face of thecompressor housing 207. A second diffuser wall 413 is formed by anexternal face of the compressor-end portion of the bearing housing. Thefirst and second diffuser walls form the diffuser passage 311.

The turbine-end portion 401 of the bearing housing 209 is affixeddirectly adjoining the turbine housing 205 using connectors thatcompress a turbine housing flange 421 against a first bearing-housingflange 423 on the turbine-end portion. The central-portion 403 of thebearing housing is affixed directly adjoining the compressor housing 207using connectors that compress a compressor-housing flange 425 against asecond bearing-housing flange 427 of the central portion. Thecompressor-end portion 405 of the bearing housing forms a circular bodythat axially is conformingly received within the compressor housing toseparate the second inlet passage 307 from the diffuser passage 311.

As is typical for single-sided radial compressors, the compressorhousing 207 includes a portion that substantially conforms to the spacethrough which the first set of impeller blades 301 is configured torotate, leaving only a small gap there-between. This portion thus formsa first shroud wall 429 for the first set of impeller blades.

Axially adjacent to the first shroud wall 429, the compressor housing207 forms a first inlet wall 431 that leads to a first inducer formed byleading edges of the first set of impeller blades 301. At a firstexducer formed by trailing edges of the first set of impeller blades,the first shroud wall adjoins the first diffuser wall 411. Thus, thecompressor housing serially forms the first inlet wall, the first shroudwall, and the first diffuser wall.

The compressor-end portion 405 of the bearing housing 209 extends aroundthe second set of impeller blades 303. A shroud portion of thecompressor-end portion substantially conforms to the space through whichthe second set of blades is configured to rotate, leaving only a smallgap there-between. This shroud portion thus forms a second shroud wall441 for the second set of impeller blades.

Axially adjacent to the second shroud wall 441, the compressor-endportion 405 forms a second inlet wall 443 that leads to a second inducerformed by leading edges of the second set of impeller blades 303. Thesecond inlet wall is are in turn adjacent to the second wall 409 of thesecond inlet passage 307. At a second exducer formed by trailing edgesof the second set of blades, the second shroud wall adjoins the seconddiffuser wall 413. Thus, the compressor-end portion 405 serially formsthe second wall of the second inlet passage, the second inlet wall, thesecond shroud wall, and the second diffuser wall.

Synergies

The configuration of the present embodiment is significant for a numberof reasons, and it is particularly effective for overcoming theefficiency limitations that limit the effectiveness of turbochargers onsmall gasoline powered engines, where the practical limitations ofconventional axial turbines render them relatively ineffective forpractical and efficient use.

The present invention provides an effective turbine with large bladesthat can be efficiently manufactured, even in small sizes. Thecomparatively large size and small number of axial turbine blades arewell suited to casting in small sizes when smaller blades might be toosmall for conventional casting techniques. The high speed flow and largeblades do not require manufacturing tolerances that may be limiting whenapplied to a very small turbine.

Singularly, the use of either a no-axial-load turbine or a no-axial loadcompressor is less efficient than their conventional axially loadedcounterpart. Moreover, turbines and compressors are typically configuredto have partially offsetting axial loads. Although these loads are farfrom perfectly matched, they do provide at least some relief from axialloads. If only one component (i.e., either the turbine or thecompressor) creates no axial load, the remaining load from the othercomponent is not partially offset, and even greater axial loads occur,requiring an even larger thrust bearing.

In the present invention, a no-axial-load compressor is combined with ano-axial-load turbine, allowing for the use of much more efficientthrust bearings. It is believed that in some embodiments the thrust loadrequirements may be as small as only 20% of the conventionalcounterparts. Bearings configured to carry such small loads can beadapted to be substantially more energy efficient. As a result, despitethe potentially lower efficiencies of some of the system components, theoverall system efficiency of the turbocharger may be significantlyhigher than in a conventional counterpart.

Other Aspects

While many conventional turbochargers are designed to produce nodownstream swirl, some embodiments of the present invention may beconfigured with blades that produce either negative or even positiveswirl. In designing a turbine under the present invention, theproduction of downstream swirl might be considered of less interest thanin the efficient extraction of energy while producing little or no axialloading.

It is to be understood that the invention comprises apparatus andmethods for designing and producing the inserts, as well as for theturbines and turbochargers themselves. Additionally, the variousembodiments of the invention can incorporate various combinations of thefeatures described above. In short, the above disclosed features can becombined in a wide variety of configurations within the anticipatedscope of the invention.

For example, while the above-described embodiment is configured as aforward-flow turbocharger (i.e., the exhaust gas stream is streamedthrough the turbine wheel so as to come axially out the end of theturbocharger), other embodiments may be configured with a reverse flowin which the exhaust gas stream passes through the turbine wheel in adirection toward the compressor. Such a configuration, while it mightnot fit in the standard spaces allotted for internal combustion engineturbochargers, exposes the bearing housing to less heat and pressure.Also, while the described embodiment uses a wheel with cantilevered(i.e., free-ended) blades that are radially surrounded by an unmovinghousing shroud, other embodiments employing a shrouded wheel (i.e., awheel having an integral shroud that surrounds the blades and rotateswith them) is within the scope of the invention.

While particular forms of the invention have been illustrated anddescribed, it will be apparent that various modifications can be madewithout departing from the spirit and scope of the invention. Thus,although the invention has been described in detail with reference onlyto the preferred embodiments, those having ordinary skill in the artwill appreciate that various modifications can be made without departingfrom the scope of the invention. Accordingly, the invention is notintended to be limited by the above discussion, and is defined withreference to the following claims.

What is claimed is:
 1. A turbocharger, comprising: a turbochargerhousing including a turbine housing, a bearing housing that is integral,and a compressor housing, wherein the turbine housing is axiallyconnected to the bearing housing, and the bearing housing is axiallyconnected to the compressor housing, and wherein the bearing housinghouses bearings; a rotor configured to rotate within the turbochargerhousing along an axial axis of rotor rotation, the rotor including aturbine wheel, a two-sided compressor wheel, and a shaft extending alongthe axis of rotor rotation and connecting the turbine wheel to thecompressor wheel, wherein the bearings constrain the rotor from movementother than axial rotation, and wherein the compressor wheel includes afirst set of compressor blades facing axially away from the turbinewheel and a second set of compressor blades facing axially toward theturbine wheel; wherein the compressor housing forms a first shroud wallthat substantially conforms to a space through which the first set ofcompressor blades is configured to rotate, leaving only a small gapthere-between; and wherein the bearing housing forms a second shroudwall that substantially conforms to a space through which the second setof compressor blades is configured to rotate, leaving only a small gapthere-between.
 2. The turbocharger of claim 1, wherein: the bearinghousing sequentially includes a turbine-end portion, a central portion,one or more supports, and a compressor-end portion; the turbine-endportion and central portion forms structures that support the bearings;the compressor end portion forms the second shroud wall; and the one ormore supports extend between the central portion and the compressor-endportion.
 3. The turbocharger of claim 2, wherein: the one or moresupports are a plurality of supports that are radially separated fromand circumferentially spaced around the axis of rotor rotation; and theplurality of supports form a plurality of radial openings into thebearing housing between the compressor-end portion and the centralportion.
 4. The turbocharger of claim 2, wherein the central-portion ofthe bearing housing is affixed directly to the compressor housing. 5.The turbocharger of claim 1, wherein: the bearing housing sequentiallyincludes a first portion, and a second portion; the first portion formsa first wall of an inlet passage configured to pass inlet air to aninlet of the second set of compressor blades; and the second portionforms a second wall of the inlet passage.
 6. A bearing housing for usewith the a rotatable two-sided compressor wheel of a turbocharger, thecompressor wheel having a first set of blades and a second set ofblades, comprising: an integral body sequentially including a first-endportion, a central portion, one or more supports, and a second-endportion; wherein the first-end portion and central portion formsstructures configured to support bearings; the second-end portion formsa shroud wall that substantially conforms to a space through which thesecond set of compressor blades is configured to rotate, leaving only asmall gap there-between; and the one or more supports extend between thecentral portion and the second-end portion.
 7. The turbocharger of claim6, wherein: the one or more supports are a plurality of supports thatare radially separated from and circumferentially spaced around the axisof rotor rotation; and the plurality of supports form a plurality ofradial openings into the bearing housing between the compressor-endportion and the central portion.
 8. The turbocharger of claim 7, whereinthe central portion forms a first wall of an inlet passage configured topass inlet air to an inlet of the second set of compressor blades, andwherein the second-end portion forms a second wall of the inlet passage.